Turbocompressor Antisurge Control by Vibration Monitoring

ABSTRACT

The proposed mechanical method of turbocompressor surge detection uses vibration signals from vibration monitoring equipment mounted on the compressor components to detect a surge event and provide antisurge control thereby. This method utilizes only mechanical information to identify surge, as compared to present day antisurge controllers that use compressor thermodynamic information such as flow, pressure, and temperature to locate a compressor&#39;s operating point on a compressor map compared to a surge region.

CROSS REFERENCE TO RELATED APPLICATIONS

Not applicable.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates generally to an automatic control scheme. More particularly the present invention relates to a method and apparatus for protecting a turbocompressor from surge by monitoring crucial vibration information and acting thereon.

2. Background Art

Compressor surge and stall represent unstable operating regimes of axial and centrifugal turbocompressors. Either mode of instability may lead to compressor damage. First, rotor vibration due to the unsteady flow of stall can cause flutter and associated fatigue. The flow reversal of surge results in an increase in temperature within the compressor. At the same time, the reversed flow and pressure variations between the intake and discharge ends of the compressor cause rapid changes in axial thrust, thereby risking damage to the thrust bearing and causing blades or vanes to rub on the compressor housing. Furthermore, abrupt speed changes may occur, possibly resulting in overspeed or underspeed of the compressor rotor.

The aerodynamics of rotating stall and surge have been investigated extensively. This research has led to common industry definitions of local stall, stage stall, stall zone, surge and rotating stall. Local stall is a flow separation or reversal in either an impeller or diffuser in a limited angular range. Stage stall is when a local stall increases to the point where one in a series of centrifugal impellers (and associated inlet and discharge stationary components) experiences reversed flow in part of its cross-sectional flow area. In any stalled condition, the overall flow is still in a forward, pressurizing direction. A stall zone is any cross-sectional area of an impeller or diffuser undergoing a flow perturbation and which manifests symptoms of a stall in a compressor. Surge is defined as the periodic oscillations of the bulk compressor flow and its associated periodic pressure swings. If these oscillations include flow reversals, deep surge is said to have occurred. Rotating stall comprises stall zones covering several blades and passages. The stall zones propagate circumferentially at a fraction of rotor speed. The number of stall zones and the propagating rates vary between compressors.

A compressor and traditional antisurge control system are shown in FIG. 1. Turbocompressor antisurge control methods of the prior art make use of thermodynamic information taken at the inlet and outlet of the compressor 100 driven by a driver 105. This information typically comprises at least a pressure differential signal gleaned from an obstruction flow meter 110 and transmitted by a flow transmitter 115, suction pressure, transmitted by a suction pressure transmitter 120, and discharge pressure, transmitted by a discharge pressure transmitter 130. These signals are fed into an antisurge controller 140 where the signals are analyzed and a closed loop response is calculated. This closed loop response determines the set point of an antisurge valve 150. Signals representing other thermodynamic data, such as one or more temperatures, may also be used by the antisurge controller. Mechanical parameters such as compressor rotational speed, inlet guide vane position, or discharge guide vane position may also be measured and transmitted to the antisurge control system. Typical compressor drivers 105 comprise steam or gas turbines and electric motors.

The antisurge valve 150 may be a recycle valve such as that shown in FIG. 1, or a blowoff valve 250, as illustrated in FIG. 2 and as might be used for air, nitrogen, and sometimes CO₂ compressors. Surge is avoided or recovered from by increasing the flow rate through the compressor via opening the antisurge valve.

Traditionally independent of the antisurge control system, vibration data are taken at radial bearing locations, at the thrust bearing, and at other locations on the compressor to monitor movement and vibration of the compressor rotor or impeller shaft. During operation, the compressor shaft is held against a thrust bearing with slight movement depending on operating rotational speed and other conditions.

Compressor surge is described in several textbooks and many articles. One such textbook is Aircraft Propulsion by Saeed Farokhi (ISBN 978-0-470-03906), published by John Wiley and Sons, is hereby incorporated in its entirety by reference. Simply speaking, surge can be defined as a point where the compressor can no longer maintain an adequate pressure difference to continue forward flow, and a bulk flow reversal occurs. Detecting the rapid reversal in flow when using an obstruction flow meter is fairly straightforward. Some compressors, however, are not fitted with a flow meter of any kind. Further, sensors and transmitters can fail so other methods to detect surge and to provide antisurge control would be of value.

There is, therefore, a need for a method and apparatus to detect surge and protect a compressor therefrom using signals other than those of thermodynamic nature.

BRIEF SUMMARY OF THE INVENTION

An object of the present invention is to provide a method and apparatus for effectively detecting turbocompressor surge using mechanical signals such as displacement and vibration. Another object of this invention is to provide antisurge control based on information gleaned from these mechanical signals.

As explained above, compressor surge is the bulk reversal of the flow direction in the compressor. That is, effectively across the full span and effectively around the entire annulus of the compressor, the fluid flow direction reverses, or at least drops to a very low level, because the compressor cannot maintain an adequate differential pressure from inlet to outlet. Because the differential pressure across the compressor rotor drops, and because of the reversed flow direction, the compressor's rotor shaft is free to move off its thrust bearing. Significant axial displacement and vibration results.

The present method monitors the axial position of the rotor shaft and detects the event when the rotor shaft is displaced from the thrust bearing, indicating surge. The axial vibration amplitude, measured in a specific frequency band, increases during surge.

When surge is detected using this method, a control system can take corrective action to extract the compressor from its surge condition.

The use of vibration data to detect surge provides an independent method to the flow path (thermodynamic) method used in the common, present day, antisurge control system. Thermodynamic transmitter failures or compressor map configuration errors do not affect the ability to detect surge based on vibration data.

The proposed system uses narrow-band frequency analysis on the vibration data. Vibration monitoring and Fourier transforms are outlined in many undergraduate textbooks. Two such textbooks are Fundamentals of Mechanical Vibrations by S. G. Kelly (2000) (ISBN 0-70-230092-2), published by McGraw-Hill, and Elements of Vibration Analysis by L. Merovitch (1975) (ISBN 0-70-041340-1), published by McGraw-Hill. Both of these texts are hereby incorporated in their entirety by reference.

The technique of the present invention allows the system to monitor specific frequency bands that tend to only respond to surge. These bands can be set for each controller based on the ambient noise floor and result in significantly higher signal-to-noise ratio when surge occurs.

Surge detection is based on a ratio of the current vibration to background vibration in each band. The background is continuously monitored and a mean value is calculated over a set time period. A set-point is calculated based on a ratio of the current value and the background value, helping to improve the signal to noise ratio. The magnitude of this set-point is determined based on the type compressor and operating conditions and may be customized for each compressor. When the set-point has been exceeded, surge is imminent or has occurred. This is true for both the axial and radial vibration.

In addition to vibration, the position of the rotor is also monitored. Axial position indicates movement of the rotor from the thrust bearing. During operation, the differential pressure across the rotor causes a small axial displacement of the rotor. When this displacement exceeds an allowable limit, surge has occurred.

Typically eddy current proximity probes are used to measure the vibration and position of the compressor rotor. Sensors other than axial and radial proximity may also be used instead of or in addition to proximity sensors to detect surge if they are mounted in a suitable location and the background noise is sufficiently low.

The novel features believed to be characteristic of this invention, both as to its organization and method of operation together with further objectives and advantages thereto, will be better understood from the following description considered in connection with the accompanying drawings in which a presently preferred embodiment of the invention is illustrated by way of example. It is to be expressly understood however, that the drawings and examples are for the purpose of illustration and description only, and not intended in any way as a definition of the limits of the invention.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

FIG. 1 is a schematic of a compressor and antisurge control system of the prior art;

FIG. 2 is a schematic of a compressor and vibration monitoring system;

FIG. 3 is a schematic of the antisurge control system and the vibration monitoring system;

FIG. 4 a is a logic diagram of a scheme for detecting surge using displacement data;

FIG. 4 b is a logic diagram illustrating the calculation of background displacement;

FIG. 4 c is a logic diagram of a scheme for detecting surge using vibration data;

FIG. 4 d is a logic diagram illustrating the calculation of background vibration;

FIG. 5 is a representative compressor map showing a surge limit and a surge control curve;

FIG. 6 is a representative compressor map showing an original and a new surge limit and an original and a new surge control curve;

FIG. 7 a is a logic diagram showing the calculation of a process variable for a PID loop from displacement data;

FIG. 7 b is a logic diagram showing the calculation of a process variable for a PID loop from vibration data;

FIG. 8 is a frequency plot of compressor rotor shaft axial displacement;

FIG. 9 is a frequency plot of compressor rotor shaft radial displacement; and

FIG. 10 illustrates a vibration sensor affixed externally to the compressor.

DETAILED DESCRIPTION OF THE INVENTION

The compressor 100 is equipped with a vibration monitoring system, including a vibration monitor 200 and one or more vibration sensors 210, 220, such as an axial displacement, velocity, or acceleration sensor 210, and radial displacement, velocity, or acceleration sensors 220. The vibration monitor 200 provides signal conditioning for the purpose of more accurately detecting surge. Additionally, the vibration monitor provides a signal that may be conveyed to an antisurge controller 140, or directly as a set point to the antisurge valve 150, 250 to avoid, prevent, or recover from a compressor surge. Thus, the vibration monitor 200 may be part of a monitoring system that generates a compressor stability indication based on the mechanical measurements described above. The sensors 210, 220 may include sensors 210, 220 operatively attached to the bearings of compressor rotor shaft 230. A thrust bearing 240 as well as a plurality of radial bearings 245, are illustrated along the compressor rotor or impeller shaft 230 in FIG. 2. The thrust bearing 240 is intended to provide a variable axial force to counter a resultant force due to the pressure forces and the axial component of the substantial derivative of the momentum of the fluid through the compressor. The radial bearings 245 are intended to provide for relatively friction-free rotation of the compressor shaft 230 and to restrict radial displacement of the shaft 230 by presenting a radial-directed force countering any radial component of force presented by the rotor shaft 230.

The axial vibration sensor 210 senses axial displacement, velocity, or acceleration of the compressor shaft 230 at the thrust bearing 240. A signal representing this measurement is transmitted to the vibration monitor 200. Similarly, the radial bearings 245 are shown with radial sensors 220 operatively attached thereto. The radial displacement sensors 220 for the radial bearings 245 transmit radial shaft displacement, velocity, or acceleration signals to the vibration monitor 200. Generally, a rotational speed sensor 260 is provided to sense the compressor shaft's angular speed. The signal from the speed sensor 260 is transmitted to the vibration monitor 200. This signal may be unnecessary, especially for a constant speed driver, such as many electric motors.

Ultimately, an antisurge valve 350 must be actuated under surge conditions to increase the flow rate through the turbocompressor. The antisurge valve 350 may be a recycle valve 150 or a blowoff valve 250. On rare occasion, a compressor's purpose is to provide a vacuum, in which case the antisurge valve is disposed on the suction side of the air compressor, and is actuated the same as the blowoff valve 250. The vibration monitor 200 may provide the antisurge valve position set point directly, as indicated in FIG. 2. Alternatively, the vibration monitor 200 may provide information to the antisurge controller 140 and the antisurge controller then provides the antisurge valve position set point as shown in FIG. 3.

FIG. 4 a illustrates a logic chart where the current displacement 405 and background displacement 410 are both calculated using only frequencies within the same predetermined frequency band, as indicated. The predetermined frequency band may be a function of the compressor's rotational speed. This simple logic chart is used to determine if the compressor 100 is in surge or not. As illustrated in FIG. 4 b, the background displacement, D_(b) 410, is calculated by continuously monitoring the shaft displacement in the predetermined frequency band during a predetermined duration of time, Δt, while the compressor 100 is not in surge and a mean value, D_(b) 410, is calculated as follows in an averaging operation, D 450:

$D_{b} = {\frac{1}{\Delta \; t}{\int_{t_{0}}^{t_{0} + {\Delta \; t}}{D\ {t}}}}$

where D is a current displacement level, calculated as a suitable vector norm such as a Root Mean Squared (RMS) value of displacement. The background displacement, D_(b) 410, may be recalculated at different operating conditions any time the compressor is not in surge.

The difference between the current displacement level, D_(c) 405, and the background displacement level, D_(b) 410, is determined in a difference operation 415. In other words, d=D_(c)−D_(b). The absolute value of d is found in the absolute value operation 420, or |d|=|D_(c)−D_(b)|. The background displacement level, D_(b) 410, is divided into the absolute value of d, as:

${r = \frac{{D_{c} - D_{b}}}{D_{b}}},$

in the division operation 425. A set point, R, may be a function 427 of the background displacement level, D_(b) 410, such as (1+n)D_(b), where n is a number greater than zero. For instance, if n=0.1, when the absolute value of d exceeds the background displacement level, D_(b) 410, by 10%, then r=R.

As long as r<R, the comparator function 430 returns a false, thus concluding the compressor 100 is not in surge. When r≧R, the comparator function 430 returns a true, thus concluding the compressor 100 is in surge.

FIG. 4 c illustrates a logic chart where vibration—velocity or acceleration—is used to detect surge. The current vibration 435 and background vibration 440 are both restricted within the same predetermined frequency band, as indicated. The predetermined frequency band may be a function of the compressor rotational speed. As illustrated in FIG. 4 d, the background vibration, V_(b) 440, which may be velocity or acceleration, is calculated by continuously monitoring the shaft vibration in the predetermined frequency band during a predetermined duration of time, Δt, while the compressor 100 is not in surge and a mean value, V_(b) 440, is calculated as follows in an averaging operation, V 455:

$V_{b} = {\frac{1}{\Delta \; t}{\int\limits_{t_{0}}^{t_{0} + {\Delta \; t}}{V{t}}}}$

where V is a current vibration level, calculated as a suitable vector norm such as an RMS value of velocity or acceleration. Those of ordinary skill in the art are well aware of the calculation of an RMS value:

$V = \sqrt{\frac{1}{N}{\sum\limits_{i}^{N}\; V_{i}^{2}}}$

The background vibration, V_(b) 440, may be recalculated at different operating conditions any time the compressor is not in surge.

The ratio of the current vibration level, V_(c), 435, to the background vibration level, V_(b) 440, is determined in a division operation 445. In other words,

$r = {\frac{V_{c}}{V_{b}}.}$

A set point, R, may be a function 427 of the background vibration level, V_(b) 440, such as (1+n)V_(b), where n is a number greater than zero. For instance, if n=0.1, when the current vibration level, V_(c), 435, exceeded the background vibration level, V_(b) 440, by 10%, then r=R.

As long as r<R, the comparator function 430 returns a false, thus concluding the compressor 100 is not in surge. When r≧R, the comparator function 430 returns a true, thus concluding the compressor 100 is in surge.

The above conclusions may be used as illustrated in FIGS. 5 and 6. FIG. 5 illustrates a representative compressor performance map, commonly referred to as a compressor map. Those of ordinary skill in this art are familiar with compressor maps. The abscissa and ordinate variables are preferably dimensionless parameters or derived from dimensionless parameters obtained from similitude. The abscissa variable, q, is frequently related to the flow rate through the compressor 100. The ordinate variable, π_(c), is frequently a static pressure ratio or related to a mass specific energy added to the compressed fluid. A more complete list of possible coordinate systems may be found in U.S. Pat. No. 5,508,943, which is hereby incorporated in its entirety by reference.

The individual curves having non-positive slopes in FIG. 5 are performance curves at different compressor rotational speeds. Each curve is for a different value of corrected speed, N_(c), which is a function of the compressor rotational speed, N.

The left-most curve 500 is the surge limit curve, surge limit line, or simply surge limit. To the left of and above the surge limit 500, the compressor's operation is unstable, and is characterized by periodic reversals of flow direction. This is surge as defined previously. The actual surge limit is sometimes unknown or only guessed at. In this case, the best guessed location of the curve is used in designing an antisurge control system for the compressor. The other curve having a positive slope in FIG. 5 is known as the surge control curve or surge control line. It is displaced toward the stable operating region from the surge limit by a safety margin. This curve is defined by an antisurge control system designer or field engineer based on experience or tests.

A typical antisurge control system will incorporate a digital depiction of the compressor map such that the control system can compare the location of the compressor's operating point to the surge control curve.

Consider a compressor operating point 620 as illustrated in FIG. 6. The surge limit 500 is assumed correct by the antisurge control designer, but may be inaccurate for a number of reasons, including compressor performance degradation over time. The surge control curve 510 is defined based on the assumed location of the surge limit 500. If the analysis illustrated in FIG. 4 a concludes the compressor 100 is in surge when its operating point 620 is as shown in FIG. 6, the antisurge control system as schematically illustrated in FIG. 1 may use that information to automatically relocate its surge limit curve 600 and surge control curve 610 as shown.

Other uses of the conclusions drawn from the logic diagram of FIG. 4 a include initiating an alarm, either visual or audible, to notify operators of a surge condition; and initiating the recording of operating parameters, such a record being archived for analysis to determine the cause of the surge event.

FIG. 7 a illustrates another logic chart where the background displacement and current displacement are both quantified within the same predetermined frequency band, as indicated. The value of r is calculated in exactly the same fashion as detailed for FIG. 4 a:

$r = {\frac{{D_{c} - D_{b}}}{D_{b}}.}$

The set point, R, is used in a difference operation 720, to calculate the value r−R. As above, R may be a function 725 of the background displacement, D_(b), as illustrated. The value, r−R, is used in two separate branches of the logic path. In the lower branch, the absolute valve of r−R is determined in an absolute value operation 730. In the upper branch, the value r−R remains unchanged. In a summation operation 740, the sum of these two values, i.e., r−R+|r−R| is found. This value must be nonnegative. This last sum is halved in a halving operation 750 before it is used as a process variable, PV, in a Proportional, Integral, Differential (PID) loop. The PID loop then calculates the set point for the recycle valve 350.

In the PID loop, the process variable, PV, signal may be processed to, for instance, reduce noise. Then an output of the PID loop is calculated as:

$X_{sp} = {{P \cdot {PV}} + {I{\int\limits_{t}^{t + t_{l}}{({PV}){\; \tau}}}} + {D\frac{({PV})}{t}}}$

which is used as the set point for the antisurge valve. In this equation, P is the coefficient for the proportional term, I is the coefficient for the integral term, D is the coefficient for the derivative term, and t_(l) is the loop time of the control loop. Those of ordinary skill in the art are well familiar with PID loops.

FIG. 7 b illustrates still another logic chart where the background vibration and current vibration are both quantified within the same predetermined frequency band, as indicated. Here, vibration connotes velocity or acceleration. The ratio of the current vibration level, V_(c), to the background vibration level, V_(b), is determined in a division operation 710,

$r = {\frac{V_{c}}{V_{b}}.}$

The set point, R, is used in a difference operation 720, to calculate the value r−R. As above, R may be a function 725 of the background vibration, V_(b), as illustrated. The value, r−R, is used in two separate branches of the logic path. In the lower branch, the absolute value of r−R is determined in an absolute value operation 730. In the upper branch, the value r−R remains unchanged. In a summation operation 740, the sum of these two values, i.e., r−R+|r−R| is found. This value must be nonnegative. This last sum is halved in a halving operation 750 before it is used as a process variable, PV, in a Proportional, Integral, Differential (PID) loop. The PID loop then calculates the set point for the recycle valve 350.

In the PID loop, the process variable, PV, signal may be processed to, for instance, reduce noise. Then an output is calculated as above, and is used as the set point for the antisurge valve.

A plot of a Fourier transform of the axial vibration data taken from a compressor is shown in FIG. 8. The spike at about 412 Hz represents the rotational speed of the compressor. When surge occurs, the axial vibration level in a range 800 of low frequencies, i.e. 10-70 Hz increases dramatically. Hence, a band of frequencies in this range 800 proves most useful for monitoring for surge in a preferred embodiment of this invention.

The Fourier transform of radial vibration of a compressor is plotted in FIG. 9. Compressor surge will result in increased vibration in a frequency band 900 of 40-60% of the compressor's rotational speed. A band of radial frequencies in this frequency range 900 is used for monitoring for surge in a preferred embodiment of this invention. The exact band is preferably determined individually for each compressor by test.

Another preferred embodiment of the present invention is shown in FIG. 10. This embodiment is particularly suited to retrofits, where a compressor 100 was not originally outfitted with a vibration monitoring system. A sensor 1010, which may sense position, velocity, or acceleration, is installed on an external component of the compressor 100, such as the housing, volute, piping, etc. where it will provide accurate measurements of the vibrations caused by compressor surge. The same calculation methods and application of the results as explained and illustrated herein are used with the data gathered from such an externally mounted sensor. The location of the externally mounted sensor should be chosen to minimize background noise.

The above embodiments are the preferred embodiment, but this invention is not limited thereto, nor to the figures and examples given above. It is, therefore, apparent that many modifications and variations of the present invention are possible in light of the above teachings. It is, therefore, to be understood that within the scope of the appended claims, the invention may be practiced otherwise than as specifically described. 

1. A method of antisurge control for a turbocompressor instrumented with at least one vibration sensor, and an antisurge valve, the method comprising: (a) selecting a frequency band; (b) acquiring a first set of vibration data from the at least one vibration sensor when the turbocompressor is not in surge; (c) calculating a background vibration level based on the first set of vibration data in the frequency band; (d) acquiring a second set of vibration data from the at least one vibration sensor whether the turbocompressor is in surge or not; (e) calculating a current vibration level based on the second set of vibration data in the frequency band; (f) calculating a process variable as a function of the background vibration level and the current vibration level; (g) selecting a set point, R; (h) comparing the process variable to the set point, R; and (i) using a comparison of the process variable to the set point, R, in a turbocompressor antisurge control system.
 2. The method of claim 1 wherein using the comparison of the process variable to the set point in a turbocompressor antisurge control system comprises opening the antisurge valve based on the comparison of the process variable to the set point.
 3. The method of claim 1 wherein using the comparison of the process variable to the set point in a turbocompressor antisurge control system comprises: (a) determining if the turbocompressor is in surge based on the comparison of the process variable to the set point; (b) comparing a turbocompressor operating point to a surge control curve on a turbocompressor performance map; and (c) if the turbocompressor is in surge, moving the turbocompressor surge control curve based on the comparison of the operating point to the surge control curve.
 4. The method of claim 1 wherein the frequency band is selected based on tests run on the turbocompressor.
 5. The method of claim 1 wherein a vibration data type is selected from the group consisting of displacement, vibration, and acceleration.
 6. The method of claim 1 wherein calculating a background vibration level comprises: (a) repeatedly transforming the acquired vibration data using a fast Fourier transform over a predetermined period of time; (b) repeatedly calculating a vector norm for the acquired vibration data within the frequency band over the predetermined period of time; and (c) calculating an average vector norm based on the vector norm values calculated repeatedly over the predetermined period of time.
 7. The method of claim 6 wherein the vector norm comprises a root mean square value.
 8. The method of claim 1 wherein calculating a current vibration level comprises: (a) transforming the acquired vibration data using a fast Fourier transform; and (b) calculating a vector norm for the acquired vibration data within the frequency band.
 9. The method of claim 8 wherein the vector norm comprises a root mean square value.
 10. The method of claim 1 wherein the vibration data comprises displacement data, and wherein calculating a process variable comprises: (a) calculating a difference, d, between the current displacement level and the background displacement level; (b) calculating an absolute value of d; and (c) calculating a ratio, r, of the absolute value of d and the background displacement level;
 11. The method of claim 1 wherein the vibration data type is chosen from a group consisting of velocity data and acceleration data, and wherein calculating a process variable comprises calculating a ratio, r, of the current vibration level to the background vibration level.
 12. The method of claim 1 wherein selecting a set point comprises selecting a set point as a function of the background vibration level.
 13. The method of claim 12 wherein the function of the background vibration level comprises a linear function.
 14. The method of claim 10 wherein comparing the process variable to the set point comprises: (a) computing a difference, r−R, by subtracting the set point, R, from the ratio, r, of the absolute value of d and the background vibration level; and (b) calculating a second process variable by subtracting an absolute value of the difference, r−R, from the value of the difference, r−R.
 15. The method of claim 11 wherein comparing the process variable to the set point comprises: (a) computing a difference, r−R, by subtracting the set point, R, from the ratio, r, of the current vibration level to the background vibration level; and (b) calculating a second process variable by subtracting an absolute value of the difference, r−R, from the value of the difference, r−R.
 16. The method of claim 14 additionally comprising: (a) passing the second process variable into a PID loop as a PID loop process variable; (b) calculating an antisurge valve set point within the PID loop; and (c) manipulating a position of the antisurge valve based on the antisurge valve set point.
 17. The method of claim 15 additionally comprising: (a) passing the second process variable into a PID loop as a process variable; (b) calculating an antisurge valve set point within the PID loop; and (c) manipulating a position of the antisurge valve based on the antisurge valve set point.
 18. The method of claim 10 wherein comparing the process variable to the set point comprises determining if the process variable is greater than the set point.
 19. The method of claim 11 wherein comparing the process variable to the set point comprises determining if the process variable is greater than the set point.
 20. The method of claim 1 wherein acquiring vibration data from the at least one vibration sensor comprises acquiring data selected from the group consisting of radial shaft displacement, radial shaft velocity, radial shaft acceleration, axial shaft displacement, axial shaft velocity, axial shaft acceleration.
 21. The method of claim 1 wherein acquiring vibration data from the at least one vibration sensor comprises acquiring data from the at least one vibration sensor operatively affixed to an exterior of the turbocompressor.
 22. The method of claim 1 wherein the selected frequency band is a function of a rotational speed of the turbocompressor.
 23. An apparatus for providing turbocompressor antisurge control, the apparatus comprising: (a) a turbocompressor; (b) at least one vibration sensor selected from the group consisting of a proximity probe, a velocimeter, and an accelerometer; (c) a predetermined frequency band; (d) a Fourier transform calculating function to calculate a background vibration level when the turbocompressor is not in surge, and a current vibration level, both the background and the current vibration levels calculated within the predetermined frequency band and based on the data acquired from the at least one vibration sensor; (e) a process variable calculating function to calculate a process variable as a function of the background vibration level and the current vibration level; (f) a set point, R; (g) a comparison function to compare the process variable to the set point, R; and (h) a turbocompressor antisurge control system that uses a result of the comparison function.
 24. The apparatus of claim 23 wherein the at least one vibration sensor is disposed to sense vibration selected from the group consisting of axial shaft vibration and radial shaft vibration.
 25. The apparatus of claim 23 wherein the turbocompressor antisurge control system comprises: (a) a PID loop; (b) an input that receives a process variable based on the result of the comparison function; and (c) an output operatively connected to an antisurge valve by which an antisurge valve set point is transmitted to the antisurge valve.
 26. The apparatus of claim 23 wherein the turbocompressor antisurge control system comprises: (a) a digital depiction of a turbocompressor surge control curve; (b) an operating point calculating function to calculate a turbocompressor operating point; (c) a second comparison function to compare the turbocompressor operating point to the turbocompressor surge control curve; and (d) a turbocompressor surge control curve recalculation function to move the turbocompressor surge control curve based on the comparison of the turbocompressor operating point to the turbocompressor surge control curve. 